JP2016526637A - Air handling control for opposed piston engine with uniflow scavenging - Google Patents

Abstract

In an air handling system for a uniflow scavenging, two-stroke cycle opposed piston engine, one or more engine operating condition parameters are sensed and in response to the second parameter, the engine cylinders are closed when the last port of the engine operating cycle is closed. Numerical values of air handling parameters based on the state captured within are determined, these numerical values are evaluated, and one or more of these numerical values are adjusted in response to the evaluation. Using the adjusted numerical values, the air supply air flow and the EGR flow are controlled in the air handling system.

Description

[Related applications]
This application includes subject matter related to the subject matter of the following commonly assigned applications: US Patent Application No. 13 / 068,679, published as US Patent Application Publication No. 2011/0289916, International Publication No. 2013/126347. No. 13 / 782,802 published as US 2013/0267737, published as US 2013/0174548.

  The field of the invention relates to two-stroke cycle internal combustion engines. In particular, the field of the invention relates to a uniflow scavenging opposed piston engine with an air handling system that provides pressurized charge for combustion and processes combustion products. In some embodiments, such an air handling system reduces the combustion temperature by recirculating the exhaust and mixing with the pressurized charge.

  A two-stroke cycle engine is an internal combustion engine in which a power cycle is completed by one complete rotation of a crankshaft and two strokes of a piston connected to the crankshaft. As an example of a two-stroke cycle engine, there is an opposed piston engine in which two pistons are arranged oppositely in a cylinder bore so as to reciprocate in opposite directions. The cylinder is provided with an intake port and an exhaust port which are arranged in the vicinity of both ends of the cylinder and are spaced apart in the longitudinal direction. Each of the opposed pistons opens one port when moving to the bottom center (BC) position and closes when moving from the BC to the top center (TC) position to control one of the ports. Since one of the ports provides a passage for exhausting combustion products from the bore and the other serves to allow the intake air to flow into the bore, these are referred to as “exhaust” and “intake” ports, respectively. In a uniflow scavenging type opposed piston engine, supply air flows into the cylinder from its intake port and exhaust gas is exhausted from the exhaust port, so that gas flows in one direction from the intake port to the exhaust port ("uniflow"). ).

  In FIG. 1, a uniflow scavenging two-stroke cycle internal combustion engine is embodied by an opposed piston engine 49 provided with at least a one-port type cylinder 50. For example, the engine may be a 1 port type cylinder, a 2 port type cylinder, a 3 port type cylinder, a 4 port or more port type cylinder. Each port type cylinder 50 has a bore 52 and longitudinally spaced exhaust ports 54 and intake ports 56 formed or machined at each end of the cylinder wall. Each of the exhaust port 54 and the intake port 56 includes one or more circumferential rows of openings in which adjacent openings are separated by a solid bridging portion. In some descriptions, each opening is referred to as a “port”, but the structure of the circumferential array of such “ports” is not different from the structure of the ports shown in FIG. In the illustrated embodiment, the engine 49 further includes two crankshafts 71, 72. In the bore 52, an exhaust piston 60 and an intake piston 62 are slidably disposed with their end faces 61 and 63 facing each other. The exhaust piston 60 is coupled to the crankshaft 71, and the intake piston is coupled to the crankshaft 72.

  A combustion chamber is defined in the bore 52 between the end faces 61 and 63 of the pistons when the pistons 60 and 62 are near TC. Fuel is directly injected into the fuel chamber via at least one fuel injector nozzle 100 positioned in an opening that penetrates the side wall of the cylinder 50. The fuel is mixed with the intake air that has entered the bore via the intake port 56. When the air / fuel mixture is compressed between the end faces, it reaches a temperature at which combustion occurs.

  With further reference to FIG. 1, the engine 49 includes an air handling system 51 that handles the conveyance of the supply air provided to the engine 49 and the exhaust of the gas produced by the engine 49. A typical structure for an air handling system includes an air supply subsystem and an exhaust subsystem. In the air handling system 51, the air supply subsystem is coupled to an air supply source that receives and processes fresh air to produce the air supply, and the air supply to the at least one intake port of the engine. At least coupled to the air supply channel for receiving and cooling the air supply channel carrying and the air supply (or a mixture of gases containing the supply air) prior to delivery to one or more intake ports of the engine One air cooler. Such coolers can comprise air-liquid devices and / or air-air devices, or other cooling devices. The exhaust subsystem includes an exhaust channel for conveying exhaust products from the engine exhaust port for delivery to another exhaust component.

  Still referring to FIG. 1, the air handling system 51 includes a turbocharger 120 provided with a turbine 121 and a compressor 122 rotating on a common shaft 123. Turbine 121 is coupled to the exhaust subsystem and compressor 122 is coupled to the charge subsystem. The turbocharger 120 exits from the exhaust port 54 and directly flows into the exhaust channel 124 from the exhaust port 54 or is exhausted from the exhaust manifold 125 that collects exhaust gas output through the exhaust port 54. Energy is extracted from the exhaust flowing into the channel 124. In this connection, the turbine 121 is rotated by the exhaust passing through it. As a result, the compressor 122 is rotated, whereby fresh air is compressed and air supply is generated. The air supply subsystem includes a supercharger 110. The supply air output by the compressor 122 flows through the supply air channel 126 to the cooler 127 and is sent from there to the intake port by the supercharger 110. The supply air compressed by the supercharger 110 can be output to the intake manifold 130 via the cooler 129. The intake port 56 receives the supply air sent through the intake manifold 130 by the supercharger 110. In the multi-cylinder opposed piston engine, the intake manifold 130 is preferably composed of an intake plenum communicating with the intake ports 56 of all the cylinders 50.

  In some aspects, the air handling system shown in FIG. 1 can be configured to reduce NOx (nitrogen oxide) emissions caused by recirculating and burning the exhaust in a ported cylinder of the engine. The recirculated exhaust is mixed with the charge air to reduce the peak combustion temperature, thereby reducing NOx production. This process is called exhaust gas recirculation (“EGR”). The illustrated EGR structure takes a portion of the exhaust flowing from the port 54 during scavenging and passes it through an EGR loop external to the cylinders for the flow of fresh intake air entering the supply subsystem. Carry in. The EGR loop preferably includes an EGR channel 131. The recirculated exhaust flows through the EGR channel 131 under the control of a valve 138 (this valve is also referred to as an “EGR valve”).

  In many two-stroke engines, combustion and EGR operations are monitored and optimized based on various measurements related to the amount of charge delivered to the engine. For example, the ratio of the charge air mass delivered to the cylinder ("lambda") to the reference charge mass required for stoichiometric combustion in the cylinder is used to control NOx over a range of engine operating conditions. Is done. However, in a two-stroke cycle opposed piston engine with uniflow scavenging, the number of port opening times overlaps for a part of each cycle, and some of the supply air delivered to the cylinder via the intake port is closed. Before flowing out of the cylinder. The supply air flowing out of the exhaust port during scavenging cannot be used for combustion. Therefore, the lambda value (“delivered lambda”) based on the supply air delivered to the intake port of the cylinder of the opposed piston engine with uniflow scavenging exceeds the supply air amount actually available for combustion.

  Therefore, it is necessary to improve the accuracy of air handling control in a uniflow scavenging type opposed piston engine.

  In a two-stroke cycle opposed piston engine with uniflow scavenging, the lambda is estimated or calculated based on the charge collected in the cylinder by closing the last port. In this connection, the last port to close may be either an intake port or an exhaust port. In this connection, the charge mass trapped in the cylinder due to the closing of the last port (hereinafter “last port closing” or “LPC”) is required for stoichiometric combustion in the cylinder. The ratio to the reference charge mass is called “collected lambda”. Since it is the collected charge air that is available for combustion, the collected lambda value represents the engine's potential for combustion and emissions more accurately than the delivered lambda value.

  In some aspects, combustion and EGR operation in a two-stroke cycle opposed piston engine with uniflow scavenging is monitored and controlled based on the collected state at LPC during the most recent engine operating conditions.

  In some aspects, the control of the air handling system of a two-stroke cycle opposed piston engine with uniflow scavenging is based on the charge collected in the cylinder by the last port closure.

  A method of controlling an air handling system of a uniflow scavenging, two-stroke cycle opposed piston engine is based on sensing one or more engine operating state variables and in response to the sensing step and collected during LPC. Determining values for the air handling parameters, evaluating these values, and adjusting one or more of these values in response to the evaluation step. In some aspects, the air handling parameter includes the collected lambda.

  In some further aspects, the air handling parameter is defined as the ratio of the mass of gas generated as a result of combustion to the total mass of gas collected in the cylinder after combustion prior to the start of scavenging. In addition, a burned gas fraction is also included.

  A method for controlling an air handling system of a uniflow scavenging two-stroke cycle opposed piston engine senses one or more engine operating state variables and collects them in the engine cylinder during LPC in response to this sensing step. Determining values of air handling parameters based on the determined state, evaluating these values, and adjusting the operation of the air handling system apparatus in response to the evaluation step. In some aspects, the air handling parameter includes the collected lambda. In some other aspects, the air handling parameter further includes a collected burned gas fraction.

FIG. 2 is a diagram of an opposed piston engine equipped with an air handling system with EGR, accurately labeled “prior art”.

1 is a schematic diagram showing a control mechanization section for regulating an air handling system in an opposed piston engine.

It is a control flowchart which shows the loop for calculating | requiring the numerical value of an air handling control parameter during an engine driving | operation.

It is a control flowchart which shows the process of evaluating and adjusting the numerical value of an air handling control parameter.

It is a basic diagram which shows the control mechanization part which implement | achieves the evaluation process and adjustment process of FIG.

  In order to maintain optimal control of combustion and emissions in response to changes in engine operating conditions, it is desirable to control the flow of charge air through the charge channel of a two-stroke cycle opposed piston engine with uniflow scavenging. Using the engine of FIG. 1 as a reference, FIG. 2 shows the control mechanization based on the improvements and additions useful for controlling the air handling system according to the present description of such an opposed piston engine.

  One example of a specific EGR loop structure for a two-stroke cycle opposed piston engine with uniflow scavenging is the high pressure configuration shown in FIG. 2 (not intended to be limiting). In this regard, the high pressure EGR loop circulates exhaust from a source upstream of the input to the turbine 121 to a mixing point downstream of the output of the compressor 122. In this EGR loop, the EGR channel 131 and the EGR valve 138 divert part of the exhaust from the exhaust channel 124 into the air supply channel 126, where part of this exhaust is compressed by the compressor 122. Mixed with fresh air. The actuator 141 controls the operation of the valve 138 according to the EGR control signal. When exhaust / air mixing is not required, the valve 138 is completely closed and the supply air without exhaust components is delivered to the cylinder. As the opening of the valve 138 increases, the amount of exhaust gas mixed during the supply air increases. On the other hand, as the valve 138 is gradually closed from the open state, the amount of exhaust gas mixed in the supply air decreases. This loop exposes the recirculated exhaust to the cooling effect of the two coolers 127,129. If there is an advantage with less cooling, the exhaust part can be diverted around the cooler 127 to the input of the supercharger 110, and in this alternative, the exhaust part is only supplied by the charge air cooler 129. Allow to cool. A dedicated EGR cooler that cools only the exhaust can be incorporated into the EGR channel 131 in alignment with the valve 138 or in alignment with the output port of the valve 138 and the input port to the supercharger 110.

  As shown in FIG. 2, in many aspects, the supercharger 110 is coupled to and driven by a drive mechanism 111 to the crankshaft. The drive mechanism 111 may comprise a stepped variable transmission or continuously variable transmission (CVT) device, in which case the feed mechanism 111 changes the speed of the supercharger 110 in response to a speed control signal provided to the drive mechanism. You can change the flow of mind. Alternatively, the drive mechanism 111 may be a fixed gear device. In this case, the supercharger 110 is continuously driven at a fixed speed. In such a case, the charge flow is varied by the method of the shunt channel 112 that couples the output of the supercharger 110 to its input. When the bypass valve 139 is provided in the diversion channel 112, the flow of the supply air can be changed by the modulation of the supply air pressure downstream of the supercharger output section. In some aspects, valve 139 is operated by actuator 140 in response to the shunt control signal.

  As seen in FIG. 2, the mechanization of control of the air handling system of a two-stroke cycle opposed piston engine with uniflow scavenging involves. The ECU 149 uses a valve 138, 139 (and possibly other valves), a multi-speed device, or a variable-speed device to supply the air flow and the displacement mixed with the pressurized air in response to a specific engine operating state. In this case, it is preferable that the supercharger 110 is controlled, and the variable geometry device is controlled by automatically operating the turbocharger. Of course, the operation of the valves and associated elements used in EGR can include any one or more of electrical, pneumatic, mechanical and hydraulic actuation operations. For quick and precise automatic operation, the valve is preferably a high speed, computer controlled device with a continuously variable setting. Each valve is provided with a state of being opened to allow gas to flow in (according to a setting controlled by the ECU 149) and a state of being closed to prevent the inflow of gas.

A method for controlling the air handling system of a two-stroke cycle opposed piston engine (hereinafter “engine”) with uniflow scavenging calculates the size and ratio of the combustion elements captured in the cylinder of the engine by closing the last port of the cylinder Or use various parameters to estimate. In this regard, a “combustion element” includes one or both of combustion components and combustion products. To better understand these methods, a number of air handling parameters used to represent these elements will be described with reference to the various elements of mechanization of air handling control in FIG. Unless otherwise noted, all air handling parameters in the following description have SI units.

[Air handling parameters]
W _air = mass flow rate of fresh air (kg / sec)
W _egr = mass flow rate of EGR gas (kg / sec)
W _SC = mass flow rate of supply air delivered to the cylinder (kg / sec)
W _f = commanded engine fuel injection rate (kg / sec)
M _res = mass of residue in cylinder (kg)
Mass of the gas trapped in the cylinder at Mtr = LPC (kg)
Mret = mass of delivered charge held in cylinder (kg)
M _del = mass of supply air delivered to the cylinder (kg)
m _{O2_air} = Fresh _{O 2} in air mass fraction _{m O2_egr} = mass fraction of _{O 2} in the mass fraction of _{O 2 m O2_res} = cylinder residue in in EGR _{m O2_im} = intake manifold of _{O 2} Mass fraction T _{comp_out} = Compressor discharge temperature (K)
T _egr = EGR temperature after leaving the cooler (K)
The temperature of the supply air trapped in the cylinder at T _tr = LPC (K)
[O ₂ ] _jm = proportion of volume concentration of O ₂ in the intake manifold [O ₂ ] _egr = proportion of volume concentration of O _{2 in} the exhaust [O ₂ ] _air = volume concentration of O _{2 in} fresh air Ratio (O ₂ / F) _S = Stoichiometric oxygen to fuel ratio (A / F) _S = Stoichiometric air to fuel ratio γ = Specific heat ratio N = Number of cylinders V _d = Per cylinder Replacement volume (m ³ )
Replacement volume for each cylinder at V _tr = LPC (m ³ )
R = gas constant of air (J / Kg / k)
R _O2 = oxygen gas constant (J / Kg / k)
AFR _s = Diesel stoichiometric air fuel ratio AFR _g = Overall air fuel ratio (fresh air to fuel ratio)
P _exh = exhaust pressure P _im = intake manifold pressure P _rail = fuel rail pressure Inj_time = injection timing

Under steady state conditions, the charge mass flow rate (W _sc ) can be calculated as:

W _sc = W _air + W _egr Formula 1

For engines that are not equipped with EGR, the charge mass flow is:

W _sc = W _air Formula 2

  The charge mass flow, air mass flow, and EGR mass flow are determined by applying a number of approaches to the engine.

[Measurement / estimation of air mass flow]

  One solution that is readily available for measuring fresh air mass flow is the use of a hot wire mass flow sensor 202. In this case, the sensor 202 directly outputs the mass flow rate of fresh air in real time. This approach is very simple but relatively expensive.

If the sensor 202 is omitted, cost and complexity can be reduced. Instead, the fresh air mass flow rate can be calculated, estimated, or estimated based on the exhaust fuel ratio measured by the engine exhaust NOx (or O 2) sensor 203.
Formula 3

(A / F) _s is a known fuel property. The ECU 149 can estimate the fuel flow rate as the commanded fuel injection amount.

Another method for estimating fresh air mass flow is:

W _air = W _sc −W _egr Formula 4

  Here, it is assumed that there is an excellent measurement / estimation (Equation 8 or 10 below) that can be used for the charge air mass flow and an excellent measurement / estimation (Equation 5 or 6 below) that can be used for the EGR mass flow. To do.

[Measurement / Estimation of EGR Mass Flow]

One approach to measuring EGR flow is to position the venturi in the EGR conduit 131 before the exhaust in the EGR conduit 131 downstream of the valve 138 mixes with fresh air in the conduit 126. In this case, the EGR flow rate is determined by the pressure drop across the venturi, upstream pressure and temperature. Some calibration is required to determine the flow coefficient and to accommodate pressure pulsations that vary with load and speed.
Formula 5

  Another method for calculating the EGR flow rate in the engine is to measure oxygen and exhaust in the intake manifold 130. Measurement of oxygen in engine exhaust gas is readily available from the NOx sensor 203. It can also be measured by setting a broadband O2 sensor in the exhaust stream 128. Intake manifold oxygen can also be measured in a similar manner using broadband O 2 sensor 205.

The intake manifold O2 sensor must compensate for intake manifold pressure. Once the intake O2 concentration and exhaust O2 concentration are obtained, the mass flow rate of EGR gas can be calculated as:
Formula 6

Equation 6 is based on the assumption that the process of EGR mixing with fresh air is adiabatic and isobaric (in this case, the exhaust pressure of the compressor 122), and the gas characteristics of EGR and fresh air are the same. Is. Since the O ₂ concentration in the air is a known number that changes with the atmospheric pressure and the relative humidity, it can be obtained from a map as shown in Equation 7.

It can be assumed that the EGR temperature is the same as the exhaust temperature at the turbine inlet measured by the temperature sensor 210. The EGR temperature can also be obtained from a temperature sensor (not shown) that can be installed in the EGR pipe before or behind the EGR valve 138.

[O ₂ ] _air = f (P _atm , Hum _rel ) Equation 7

[Intake mass flow estimation]

  When the mass flow rate of EGR gas and the mass flow rate of fresh air are determined, the mass flow rate of the supply air can be obtained based on Equation 1 or Equation 2 depending on the engine structure.

Alternatively, the mass flow rate of the supply air can be calculated mathematically by modeling the engine cylinder as an orifice based on the relationship shown in Formula 8.
Formula 8

Using a look-up table (LUT) indexed by engine speed (RPM), exhaust pressure (P _exh ), charge manifold pressure (P _im ), C _d (emission factor) is _calculated for each speed across the engine. Can be calibrated for / pressure ratio.

C _d = LUT (RPM, P _exh , P _im ) Equation 9

The effective orifice area (A _eff ) can be calculated based on the number of intake and exhaust ports and the port timing.

  The intake manifold pressure can be read directly from the differential intake manifold pressure sensor 207. When the sensor 207 is arranged behind the supercharger 110, the intake manifold pressure can be calculated by subtracting the estimated pressure drop of the charge air cooler 129 from the outlet pressure of the supercharger 110.

  The exhaust manifold pressure can be read directly from a pressure sensor 209 disposed within the exhaust manifold 125. In some cases, it may not be possible to place the pressure sensor directly inside the exhaust manifold. In that case, the reading value from the turbine inlet pressure sensor 210 can be used instead of the exhaust manifold pressure of Equation 8.

Another approach to estimating charge mass flow is based on pressure ratio across the supercharger 110, supercharger speed, and supercharger efficiency.

W _sc = f (sc_speed, PR _sc ) Equation 10

When the supercharger 110 is driven by the continuously variable drive unit 111, the supercharger speed (sc_speed) can be measured using a speed sensor. The supercharger pressure ratio (PR _sc ) can be calculated by dividing the intake manifold pressure by the supercharger inlet pressure. The supercharger inlet pressure can be calculated by subtracting the estimated pressure drop across the charge air cooler 127 from the pressure measured at the compressor outlet. Alternatively, the pressure sensor 211 may be arranged behind the supply air cooler 127 so that the supercharger inlet pressure can be reported directly. In this case, the compressor outlet pressure can be calculated by adding the estimated pressure drop across the charge air cooler 127 to the supercharger inlet pressure measurement.

  The pressure drop across the charge air cooler 127 can be estimated from a look-up table relating the charge air flow rate to the pressure drop. Alternatively, the pressure drop across the charge air cooler 127 can be measured directly using a differential pressure sensor.

[Estimation of air supply ratio]

When the air supply mass flow rate is obtained by any of the above methods, the air supply ratio (Λ) can be obtained. The definition of the air supply ratio of the present disclosure is shown in Formula 11.
Formula 11

[Estimation of scavenging efficiency]

Next, the scavenging efficiency is calculated. For the purposes of the air handling system control method, the scavenging efficiency is defined as:
Formula 12

The scavenging efficiency can be calculated from an empirical model that relates the charge ratio and engine speed to the scavenging efficiency. An empirical model can be developed from scavenging data collected from the engine during the mapping process.

η _sc = f (RPM, Λ ^* ) Equation 13

The air handling system control method uses the calculated scavenging efficiency to determine the engine collection efficiency defined as:
Formula 14

Combining Equation 12 and Equation 14 yields:
Formula 15

The amount of gas mass trapped in the cylinder can be calculated based on the following two-region non-isothermal model:
Formula 16

Here, ρ _del is the density of the supply air delivered to the port that is closed last, and ρ _res is the density of the residual gas in the cylinder that is closed last. These can be calculated as:
Formula 17
Formula 18a
Formula 18b

The exhaust temperature can be acquired from the sensor 210. It can also be obtained directly by placing a temperature sensor in the exhaust manifold 125. Furthermore, if the sensor is provided for measuring the cylinder pressure, Equation 17, replacing the P _im formula 18a, with the last measured at the port (P _LPC) for closing the cylinder pressure Can do.

The air mass delivered to one cylinder in each cycle can be calculated from the charge mass flow rate and engine speed as follows:
Formula 19

By replacing the values of M _tr and M _{del in} Equation 15, the collection efficiency can be calculated as shown in Equation 17:
Formula 20

[Estimation of collected lambda]

To find the collected lambda, start by calculating the mass of oxygen collected in the cylinder. In Formula 21, the mass of oxygen collected in the cylinder is obtained.
Formula 21

Where M _{O2_air} is the oxygen mass in fresh air delivered to the cylinder (Equation 22) and can be calculated as:
Formula 22
Formula 23
Formula 24

M _{O2 — egr in} the EGR delivered to the cylinder is the oxygen mass and can be calculated based on Equation 25:

Formula 25
Formula 26
Equation 27

If the EGR ratio x = (EGR / EGR + CHARGE AIR) is used, the oxygen mass in the delivered EGR can also be calculated based on the combustion stoichiometry as follows.
Formula 28

M _{O2 — res} is the mass of oxygen in the residual gas remaining in the cylinder after the scavenging process is completed. This can be calculated based on knowledge of scavenging efficiency, mass of oxygen collected from the previous cycle, and fuel injection amount, as shown in Equation 25.
Formula 29

When the intake manifold oxygen sensor is attached to the engine, the collected oxygen can be calculated as shown in Equation 30.

M _O2 — _tr = η _tr M _O2 — _im + M _{O2 — res} Equation 30

Here, M _{O2 — sc} is the total mass of oxygen delivered into the intake manifold and can be calculated as:

Formula 31
Formula 32
Formula 33

Finally, the collected lambda can be calculated as in Equation 34.

Formula 34

[Percentage of burned gas collected]

The total mass of burned gas collected in the cylinder changes with changes in engine scavenging and collection efficiency. For this parameter, burned gas is defined as the gas produced as a result of combustion (ie, CO ₂ (carbon dioxide) and H ₂ O (water)). Thus, a burned gas ratio of 1 represents stoichiometric combustion, meaning that all the oxygen in the air has been used up to convert the fuel (C _x H _y ) to CO ₂ and H ₂ O.

With the external EGR being supplied to the engine, the fraction of burned gas collected can be calculated as:

Formula 35

Here, M _egr is the mass of EGR delivered through the intake port, as shown in Equation 36.
Formula 36

The mass ratio of the burned gas in the exhaust gas (BF _exh ) can be calculated based on the combustion stoichiometry.
Formula 37

AFR _s is a known amount of diesel fuel. AFR _g is obtained by dividing the mass flow rate of fresh air by the fuel flow rate.

The mass of the residual gas is obtained from η _sc and M _tr as shown in Equation 38.

M _res = (1−η _sc ) M _tr Formula 38

The burnt gas mass fraction in the residue (BF _res ) is based on the presumed collected lambda and the collected mass.

Formula 39

Here, M _f is the mass of fuel injected into each cylinder for each cycle, and can be calculated as shown in Equation 40.

Formula 40

Equation 41 provides an alternative method for determining the fraction of burned gas that has been collected.

Formula 41

[Estimation of engine exhaust]

Once the ratio of the collected air-fuel ratio and the collected burned gas mass is estimated, an empirical model can be created to estimate the exhaust components ("engine exhaust emissions") that the engine exhausts. The major emissions of concern are engine exhaust NOx and engine exhaust soot. These can be calculated as:

_{[NO x, Soot] = f} (RPM, W f, λ tr, BF tr, P rail, Inj_time, T tr)

Here, T _tr is the temperature of the mass collected in the cylinder at the start of the cycle. This can be calculated as:
Formula 42

  The empirical model may be based on multiple look-up tables or non-parameterized mathematical functions such as neural networks and basis functions.

[Air handling control]

  Air handling control can be implemented using mechanization of air handling control based on that shown in FIG. 2, where the ECU 149 controls the operation of the air handling system by the method shown in the charts of FIGS. Can be programmed. In this regard, FIG. 3 shows a loop for determining the numerical value of the air handling control parameter based on the state collected in the cylinder of the engine in the LPC. FIG. 4 shows a process of evaluating and adjusting the numerical value of the air handling control parameter. FIG. 5 schematically illustrates a preferred control mechanization that implements the evaluation and adjustment process of FIG.

Referring to FIG. 3, the loop 300 reads available engine sensors at 302 to determine the current value of the air handling parameter at the current engine operating condition. ECU 149 uses these sensor values to determine the current engine operating conditions with respect to torque demand (load) and RPM, and a routine with a series of operations and calculations essentially corresponding to Equations 1-36. Run. In this regard, the routine first determines the fresh air value, EGR value, and air supply value at step 304, the air supply ratio at step 305, and the scavenging efficiency and collection efficiency at step 306. . Next, the routine uses the parameter values obtained in steps 304, 305, and 306 to obtain parameter values representing the collected state in steps 307 and 308. This is because these values reflect the current engine operating conditions, and the parameters associated with these parameter values are referred to as “actual”. In step 307, the actual collected mass (M _tr ) and the actual collected lambda (λ _tr ) are calculated. At step 308, the actual collected burned gas fraction (BF _tr ) and the actual collected temperature (T _tr ) are calculated.

  The actual collected parameter values determined in steps 307, 308 of loop 300 are provided to the desired collected cylinder state routine 400 shown in FIG. In some aspects, this desired collected cylinder status routine is executed by the ECU 149. This routine determines the desired collected lambda in the LPC and the desired collected burned gas percentage in the LPC based on a lookup table (map) indexed by load (engine torque) and RPM. The desired parameter values stored in the table are consistent with the desired performance and emission targets for engine operating conditions expressed in engine torque and RPM. These maps can be pre-populated with experience data based on engine dynamometer tests and stored in ECU 149 or by ECU 149.

  As shown in FIG. 4, the desired trapped cylinder status routine 400 maps to the map at step 402 to determine the desired trapped lambda and trapped burned gas percentage values for the current engine operating condition. Then, in step 404, the actual collected lambda value is evaluated by comparing it with the desired collected lambda value. This comparison step preferably includes subtracting the desired collected value from the actual collected value. If the absolute value of this difference is greater than the threshold value, then in step 405, the desired collected cylinder status routine 400 is configured to maintain fresh air flow and intake manifold pressures to keep this difference within acceptable limits. Adjust either (or both). Referring to FIG. 2, this result can be achieved by controlling a supercharger drive 111 that is presumed to be a transmission. By changing the speed of the supercharger 110, it is possible to adjust the fresh air flow or intake manifold pressure. Alternatively, when the supercharger driving unit 111 is a fixed driving device, the setting of the bypass valve 139 is changed by the bypass valve actuator 140 so that the output of the supercharger is diverted to the input unit. Results are obtained.

FIG. 5 illustrates an exemplary control mechanization that can control the actual collected lambda. The control mechanization unit includes a collected lambda control unit with a feedforward control unit 510 and a feedback control unit 512. The collected lambda feedforward controller 510 outputs a supercharger actuator set point Θ based on a map indexed by engine load and speed. In this map, experience data based on the engine dynamometer test is input in advance and stored in the ECU 149 or by the ECU 149. The collected lambda feedback controller 512 calculates the error (e _λtr ) between the desired collected lambda and the actual collected lambda, calculated by the adder 511 (step 404 in FIG. 4). Receive and convert this error into the required change in fresh air flow (ΔW _air ) or boost pressure (ΔP _im ) to minimize the error. The collected lambda feedback control unit 512 can be realized by another non-linear control unit such as a PID control unit, a gain-scheduled PID control unit, or a sliding mode control unit. The supercharger actuator controller 513 converts the changes necessary for fresh air flow or supercharging pressure into either a bypass valve position or a drive ratio change, depending on the type of supercharger drive 111 used. A supercharger actuator controller 513, which converts the change in fresh air flow (or supercharging pressure) into a supercharger drive output (ΔΘ), defines the relationship between the supercharger pressure ratio, speed and mass flow. By using together with a charger model (physical or empirical), it can be realized as a PID control unit or a PID control unit with a planned gain. At 514, the supercharger output from actuator controller 513 is added to the output of feedforward controller 510 (or subtracted according to the sign). Then, the last supercharger actuator command is sent from the ECU 149 to the supercharger actuator. Depending on the structure of the supercharger drive 111, the supercharger output command is provided as either a speed control signal to the drive 111 or a shunt control signal to the bypass valve actuator 140.

  Except that the routine uses the EGR valve 138 to change the EGR flow rate in order to minimize the error between the actual collected burned gas rate and the desired collected burned gas rate. Control of the collected burned gas ratio is performed in the same manner as in the case of collected lambda. Therefore, as shown in FIG. 4, at step 406, the desired collected cylinder state routine 400 compares the actual collected burned gas rate value with the desired collected burned gas rate value. To evaluate. This comparison step preferably includes subtracting the desired collected value from the actual collected value. If the absolute value of this difference is greater than the threshold, routine 400 adjusts the EGR flow at step 407 to keep this difference within acceptable limits. Referring to FIG. 2, this result can be achieved by controlling the setting of the EGR value 138 with the actuator 141.

FIG. 5 shows an exemplary control mechanization that can control the actual collected burned gas fraction. The control mechanization unit includes a collected burned gas ratio control unit including a feedforward control unit 520 and a feedback control unit 522. The collected burned gas ratio feedforward control unit 520 outputs an EGR value set point Θ based on a map indexed by engine load and speed. In this map, experience data based on the engine dynamometer test is input in advance and stored in the ECU 149 or by the ECU 149. The collected burned gas ratio feedback control unit 522 calculates the desired collected burned gas ratio and the actual collected burned gas ratio calculated by the adder 521 (step 406 in FIG. 4). The error in between (e _BFtr ) is received and this error is converted into the change necessary for the EGR flow (ΔW _EGR ) to minimize the error. The collected burned gas ratio feedback control unit 522 can be realized by another non-linear control unit such as a PID control unit, a gain scheduled PID control unit, or a sliding mode control unit. The EGR valve actuator control unit 523 converts a change necessary for the EGR flow into an EGR valve position. The EGR actuator control unit 523 that converts the EGR flow change into the EGR actuator output (ΔΘ) can be realized as a PID control unit or a gain-scheduled PID control unit in combination with an EGR model (physical or empirical). it can. At 524, the EGR actuator output from the actuator controller 523 is added to the output of the feedforward controller 520 (or subtracted depending on the sign). Then, the last EGR valve actuator command is sent by the ECU 149 to the EGR valve actuator. The EGR valve actuator command is provided to the EGR valve actuator 141 as an EGR control signal.

  Referring again to FIG. 4, once the error between the trapped lambda and the trapped burned gas ratio is minimized, the desired trapped cylinder state routine 400 then proceeds to step 408. Compare the actual collected temperature with a predefined value. This predefined collected temperature value is determined based on an engine dynamometer test and stored in or by ECU 149. If the actual collected temperature is found to be higher than this predefined temperature, then at step 409, routine 400 determines the desired collected lambda set point and the desired collected burned gas rate setting. Adjust the points so that the effect of the collected temperature on the discharge is minimized. Referring again to FIG. 5, adjustments to the desired collected lambda and collected burnt gas fraction are indexed at the collected temperature estimated at 502 (eg, by Equation 42). This is performed based on the tables 515 and 525. Values for these look-up tables are determined by engine dynamometer tests and stored in or by ECU 149. Next, at step 409, the collected and desired lambda set point (via 516, 511 in FIG. 5) and the collected and desired collected burned gas fraction set point (526, FIG. 5). Routine 400 readjusts the supercharger actuator output and EGR valve position to minimize the error between each.

  As the engine transitions from one operating point (engine load and speed) to another operating point (engine load and speed), the loop 300 and routine 400 described above are repeated continuously.

  The air handling control embodiments shown and described herein assume that the actual parameter values based on the state of the manifolds 125, 130 are attributed to the engine cylinders, but one or more cylinders of a production engine. Above, it should be apparent to those skilled in the art that the principles involved can be applied to the individual cylinders themselves, given the cost and space allowed to place and operate the associated sensors. Further, desirable parameter values can be obtained from empirical methods that map or synchronize these values with, for example, the number of port closures of one cylinder of a uniflow scavenging two-stroke cycle opposed piston engine operating in a dynamometer.

  Although the air handling embodiments shown and described herein have been shown and described with reference to a two-stroke cycle opposed piston engine with uniflow scavenging and EGR loop, those skilled in the art will be able to identify certain parameters. It should be understood that this is also useful in an air handling system in a two-stroke cycle opposed piston engine with uniflow scavenging but not equipped with EGR. For example, collected lambda is a useful parameter in optimizing the air handling operation to reduce the engine emissions described above. Refer to Equation 2 for this.

  Although the air handling control method has been described with reference to an opposed engine with two crankshafts, it should be understood that these structures are applicable to opposed piston engines with more than one crankshaft. is there. Furthermore, various aspects of these structures can be applied to opposed piston engines with ported cylinders that are opposed and / or located on either side of one or more crankshafts. . Accordingly, the protection afforded to these structures is limited only by the following claims.

Claims (18)

A bore (52), an axially spaced exhaust port (54), and an intake port (56) are disposed oppositely in the bore and are operable to open and close the exhaust port and the intake port during engine operation. At least one cylinder (50) provided with a pair of pistons (60, 62), an air supply channel (126) for supplying air to at least one intake port, and an exhaust for receiving exhaust from at least one exhaust port A uniflow scavenging opposed piston engine equipped with an air handling system (51), comprising a channel (124) and a supercharger (110) operating to feed air into the air supply channel,
The control mechanization unit (149) determines the value of the first collected air handling parameter (307) in response to the engine operating state, and also determines the first collected air handling parameter obtained. A uniflow scavenging oncoming piston engine, characterized in that it operates to adjust the airflow in the air supply channel (126) based on the value of.   The control mechanization unit (149) changes the speed (514) of the supercharger (110), and the first valve (139) diverts the supplied air flow from the output unit of the supercharger (110) to the input unit. The opposed piston of claim 1, wherein the opposed piston is operable to adjust a supply air flow based on the determined value of the first collected air handling parameter. engine.   The engine includes an exhaust recirculation (EGR) loop (131) provided with a loop input coupled to the exhaust channel (124) and a loop output coupled to the air supply channel (126), the control The mechanism (149) is further responsive to the engine operating condition to determine a value of the second captured air handling parameter (308), and the second captured air handling parameter. The opposed piston engine of claim 1, wherein the opposed piston engine is operable to adjust an EGR flow in the EGR loop (131) based on the determined value of. The first collected air handling parameter is the collected lambda (λ _tr ), and the second collected air handling parameter is the collected burned gas fraction (BF _tr ). Furthermore, the control mechanization unit (149)
Based on the obtained value of the collected lambda (λ _tr ), the speed (514) of the supercharger (110) is changed, and the first valve (139) is connected to the supercharger (110). ) To adjust the air supply by one of the operation of diverting from the output to the input.
The EGR flow is adjusted by operating the second valve (138) so as to increase or decrease the exhaust flow through the EGR loop (131) based on the value of the collected burned gas ratio (BF _tr ) obtained. 4. The opposed piston engine of claim 3, operable to do so. The control mechanization unit (149)
Operable to determine an actual collected lambda parameter value (502) based on current engine operating conditions;
Operable to determine a desired collected lambda parameter value (503) for the current engine operating condition;
_Operable to determine an error value (e _λtr ) based on the difference between the actual lambda parameter value and the desired lambda parameter value;
Responsive to the error value, operating to adjust the air flow (512) by changing one of the fresh air flow entering the air supply channel and changing the air supply manifold pressure. The opposed piston engine of claim 4, which is possible. The first collected air handling parameter is a collected lambda (λ _tr ), and the control mechanization unit (149)
Operable to determine an actual collected lambda parameter value (502) based on the current engine operating conditions;
Operable to determine a desired collected lambda parameter value (503) for the current engine operating condition;
_Operable to determine an error value (e _λtr ) based on the difference between the actual lambda parameter value and the desired lambda parameter value;
Responsive to the error value, operating to adjust the air supply (512) by changing one of the fresh air supply to enter the air supply channel and changing the intake manifold pressure. The opposed piston engine of claim 1, which is possible.   The control mechanization unit (149) is configured to control the flow of air supplied to the inlet of the supercharger and the supply air supplied to the intake port based on the determined value of the first collected air handling parameter. The opposed piston engine of claim 1, wherein the opposed piston engine is operable to regulate (512) the pressure of the air. The first collected air handling parameter is a collected lambda (λ _tr ), and the control mechanization unit (149)
Operable to determine an actual collected lambda parameter value (502) based on the engine operating condition;
Operable to determine a desired collected lambda parameter value (503) for the engine operating condition;
_Operable to determine an error value (e _λtr ) based on the difference between the actual lambda parameter value and the desired lambda parameter value;
Responsive to the error value, operable to adjust the speed of the supercharger (110);
The opposed piston engine of claim 7, operable to adjust a state of a valve (139) coupling the supercharger output to the supercharger input.   The EGR loop (131) further includes a valve (138) operable to regulate the flow of exhaust gas to the air supply channel (126), and the control mechanization (149) includes the second collection The opposed piston engine of claim 3, wherein the opposed piston engine is operable to adjust an exhaust flow through the EGR loop based on the determined value of a measured air handling parameter. The second collected air handling parameter is a collected burned gas ratio (BF _tr ), and the control mechanization unit (149)
Operable to determine an actual collected burned gas ratio parameter value (502) based on the engine operating condition;
Operable to determine a desired collected burned gas ratio parameter value (503) for the engine operating condition;
Based on the difference between the actual lambda burned gas ratio value and the desired lambda burned gas ratio value, and is operable to determine an error value (e _BFtr );
The opposed-piston engine of claim 9, operable to adjust the setting (523) of the valve (138) based on the error value. A method for operating the opposed piston engine of claim 1, comprising:
Generating exhaust in at least one ported cylinder (50) of the engine;
Conveying exhaust through an exhaust channel (124) from an exhaust port (54) of the ported cylinder;
Recirculating a portion of the exhaust from the exhaust channel through an EGR loop (131);
Pressurizing fresh air (122);
Mixing the recirculated exhaust from the EGR loop with the pressurized fresh air to form an air supply;
Pressurizing the air supply with a supercharger (110);
Providing the pressurized air supply to an intake port (56) of the ported cylinder;
Determining the value of the collected air handling parameters in response to engine operating conditions (307, 308);
Adjusting (512) an air supply to the intake port based on the determined value.   Adjusting the supply air flow includes changing the speed of the supercharger (110) and operating the first valve (139) to divert the supply air flow from the output part of the supercharger to the input part. 12. The method of claim 11, comprising one of   The method of claim 12, further comprising adjusting (522) an exhaust gas flow in the EGR loop (131) based on the determined value.   The method of claim 13, wherein adjusting the exhaust flow within the EGR loop comprises operating a second valve (138) to increase or decrease the exhaust flow through the EGR loop (131). A method of operating the opposed piston engine of claim 1, comprising:
Pressurizing fresh air (122) to form an air supply;
Pressurizing the air supply with the supercharger (110);
Supplying the pressurized charge to the intake manifold (130) of the engine;
In response to engine operating conditions, determining the value of the collected lambda parameter based on closure of one of the exhaust port (54) and the intake port (56) of the cylinder (50);
Adjusting (512) a supply air flow to the intake manifold based on the determined value.   The adjustment of the air supply flow includes changing the speed of the supercharger (110) and switching the first valve (139) so that the air supply flow is diverted between the output unit and the input unit of the supercharger. The method of claim 15, comprising one of operating.   The method of claim 16, further comprising adjusting (522) an exhaust flow in an EGR loop of the engine based on the determined value.   The method of claim 17, wherein adjusting the exhaust flow in the EGR loop includes operating a second valve (138) to increase or decrease the exhaust flow through the EGR loop.